Large earthmoving vehicles must not only transmit significant amounts of power through the power shift transmission to the driving wheels thereof, but also must transmit significant amounts of power to a plurality of auxiliary power units which typically are a number of hydraulic pumps. In a vehicle such as a wheel loader about half of the maximum available horsepower of the usual multicylinder diesel engine can be directed to a portion of these pumps for operation of the loader bucket. Moreover, additional power can simultaneously be directed to another portion of these pumps for operation of the steering system and other auxiliary purposes.
U.S. Pat. No. 3,586,185 issued to L. F. Clancy, et al on June 22, 1971 is illustrative of a large articulated wheel loader having a rear engine for driving a generator electrically connected to drive the wheels and for powering a plurality of hydraulic pumps located on both the rear frame and on the front frame. The bucket control system for that wheel loader includes a pair of large hydraulic pumps located on the front frame to avoid the requirement of running all of the implement hoses across the articulation joint between the main frames. In U.S. Pat. No. 3,586,185 these hydraulic pumps are mechanically driven by a front box via a front drive shaft which crosses the articulation joint. A rear gear box is driven by an undesirably elevated generator via a rear drive shaft, which rear drive shaft coaxially powers the front drive shaft.
When a mechanical drive mechanism is desired, rather than the electrical system mentioned above, the requirements change considerably. For example, where do you locate the gear train for driving the rear hydraulic pumps so that the loads therefrom are removed from the engine crankshaft bearings? Accordingly, the instant drive mechanism should provide a compact, economical and reliable drive box for these pumps that would be located in an optimum position intermediate the engine flywheel and the transmission input member, and that would not load the crankshaft bearings. Furthermore, the rear hydraulic pumps should be arranged on the drive box in a convenient and accessible location for servicing or the like, and the drive box should be compatible with another shaft system for driving the implement hydraulic pumps located on the front frame of the vehicle.
In addition to the aforementioned considerations, it was recognized that the shaft system for driving the front implement pumps would be relatively long and flexible so that a flexible coupling intermediate the engine flywheel and that front drive shaft would be required for reducing naturally excited engine pulses thereon. U.S. Pat. No. 2,961,892 issued to E. W. Spannhake on Nov. 29, 1960 recognized the need for a flexible coupling between the engine flywheel and an input shaft drivingly connected to a transmission on the one hand, and to a power take-off on the other hand. However, the flexible coupling of U.S. Pat. No. 2,961,892 is entirely unsatisfactory. That coupling includes a tubular rubber bushing and a friction type slip clutch arranged in parallel relation to the rubber bushing and axially loaded thereby. These rubber elements are unacceptable in high temperature environments because they deteriorate in use, and since their deterioration is accelerated in hot oil, special and costly provisions must be made to keep these elements dry and yet to dissipate the heat therefrom. Furthermore, in that patent the rubber bushing is ineffective when the slip clutch is holding the torque load therethrough, and torque is transmitted through the rubber bushing only when torque loads having an order of magnitude equal to or above net engine torque will cause the slip clutch to actually slip. That slip clutch setting is too high to adequately protect the components of the downstream portion of the drive mechanism.
Spring-type flexible couplings are widely used and have the advantage that they have no elastomeric elements to deteriorate. Borg-Warner Corporation of Chicago, Ill., has been active for many years in the development of spring couplings of various configurations as is exemplified by the following U.S. Pat. Nos.:
3,101,600 issued to C. V. Stromberg on Feb. 12, 1962.
3,138,011 issued to C. V. Stromberg on Aug. 17, 1962.
3,266,271 issued to C. V. Stromberg on Aug. 16, 1966.
4,139,995 issued to P. E. Lamarche on Feb. 20, 1979.
Of these, U.S. Pat. No. 3,101,600 is of interest from the standpoint that it includes two sets of spring couplings arranged in series in order to double the maximum amount of angular deflection possible without exceeding practical lengths of the compression springs.
Another reference of note is U.S. Pat. No. 4,171,627 issued to K. Fukuda on Oct. 23, 1979. It discloses a spring coupling wherein the drive torque is transmitted through a plurality of costly spring seats which support a plurality of compression springs arranged in side-by-side relation. While it can thereby handle more torque, that construction has only a single stiffness rate.
When using spring couplings it is desirable to make the spring rate thereof relatively low or soft in order to provide maximum isolation for the drive train under relatively steady state driving conditions throughout the entire operating speed range of the engine. But because of spring coil bending and various wear factors there are practical limits to the total amount of angular deflection that can be permitted, and when it is realized that torque loads above 6,779 Newton meters are sometimes experienced when making gear shifts it is further desirable to provide a relatively stiff coupling at the upper end of the torque range so as to absorb more energy prior to reaching the stage of maximum compression of the springs.
The present invention is directed to overcoming one or more of the problems as set forth above.